Screw pump rotors and ring seals for screw pump rotors

ABSTRACT

A pump rotor for a screw pump, comprising a shaft, a thread on the shaft, the thread comprising a groove disposed on an outer surface thereof, and a seal disposed in the groove. The seal and the groove are configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

BACKGROUND

The subject matter disclosed herein generally relates to screw pumps,and, more particularly, to ring seals for screw pump rotors.

In the exploration for oil and gas, the need to transport fluids (oil,water, gas, and foreign solids) from a wellhead to distant processing orstorage facilities (instead of building new facilities near thewellheads) is well understood. Twin-screw pumps are increasingly used toaid in the production of these wellhead fluids. The use of thetwin-screw pump enables increased production levels by lowering thepressure at the exit of the wellhead as well as a greater total recoveryfrom the reservoir by allowing lower final reservoir pressures beforeabandoning production.

FIG. 1 illustrates a conventional twin-screw pump 10. This figure ispresented simply to illustrate the main components of a twin-screw pumpand should not be considered as limiting the invention disclosed hereinin any way. As illustrated, the twin-screw pump 10 typically has tworotors 12 and 14 that are disposed within a rotor liner 19. Each rotorhas a shaft 18 with one or more outwardly extending sets of screwthreads 20 disposed on at least a portion of the length of the shaft 18.The shafts 18 run axially within the rotor liner 19. The two rotors 12,14 do not touch each other, but their respective opposed screw threads20 are intertwined. Pump 10 will often be driven by a motor (not shown),which motor rotates rotors 12 and 14. Typically, a drive gear 22 on oneof the shafts engages a second gear on the other shaft, such that, whenthe pump motor turns rotor 12, rotor 14 is turned at the same rate, butin an opposite direction. In operation, wellhead fluids, includingparticulate materials, are drawn into pump 10 at inlet 24. As the rotors12 and 14 are turned, rotor chambers 26 formed between adjacent threads20 displace the wellhead fluids along the rotor shafts 18 towards anoutlet chamber 28, which outlet chamber 28 is the point of greatestpressure at the center of the rotors, from where the wellhead fluids arefinally discharged from an outlet 30 of the pump 10. The rotor chambers26 are not completely sealed, but under normal operating conditions thenormal clearance spaces that exist between the rotors 12, 14 and betweeneach rotor and the rotor liner 19 are filled with transport fluid. Theliquid portion of the transport fluid in these clearance spaces servesto limit the leakage of the pumped fluids between adjacent chambers. Thequantity of fluid that escapes from the outlet side of the rotor backtoward the inlet represents the pump slip flow, which slip flow is knownto decrease the pump volumetric efficiency. As illustrated in FIG. 2 andexplained above, pump slip flow (illustrated by the arrows in FIG. 2)can occur between each rotor and the rotor liner 19. Other slip pathsinclude slip between screw tip and adjacent rotors and between faces ofthe threads 20.

Conventional twin-screw multiphase pumps currently face severalchallenges. First, assuming a fixed pressure rise per stage, as thetotal pressure rise requirement increases, the rotor length increases,resulting in an increased rotor deflection under the imposed pressureloading. This deflection creates a more eccentric alignment of therotors 12, 14 within the rotor liner 19 often resulting in excessiveslip between the rotors 12, 14 and the rotor liner 19 or contact andrubbing between the rotors 12, 14 or against the rotor liner 19.Additionally, as the pump slip flow increases, sand particulates trappedin the slip flow can lead to increased erosion or abrasion within thepump, particularly at the rotor tips by a phenomenon referred to asjetting. Such erosion or abrasion can lead to deterioration of theclearance profile and a further increase in the pump slip flow.

It would therefore be desirable to develop a pump rotor that minimizesor eliminates pump slip flow, resulting in a high differential pressureboost multiphase pump with a compact rotor length. In addition, improvedsealing between the edges of the rotor and the pump casing will alsoinsure a reduction in solid particulate erosion or abrasion withinclearances. It will also be desirable to provide a sealing system thatis durable, improves the performance of a pump and does not cause anydamage to the pump, even after the sealing system wears.

BRIEF DESCRIPTION

In accordance with one embodiment disclosed herein, a pump rotor for ascrew pump comprises a shaft, a thread on the shaft, the threadcomprising a groove disposed on an outer surface thereof, and a sealdisposed in the groove. The seal is spiraled into the thread fromstarting point of the thread at an end of the shaft. The groove and theseal are dimensioned to have a clearance between them, enabling the sealto move radially with respect to the thread as the pump rotor isdeflected. The groove and the seal form an interlocking mechanism sothat the seal is retained in the groove while allowing radialdisplacement of the seal with respect to the thread as the pump rotor isdeflected.

In accordance with another embodiment disclosed herein, a method ofreducing slip flow in a screw pump having a casing with a low-pressureinlet and a high-pressure outlet, a liner disposed inside of the casing,and a rotor disposed inside of the liner having a shaft and a threaddisposed on an outer surface of the shaft, comprises forming a groove onouter surface of the thread and disposing a ring seal in the groove suchthat the ring seal protrudes outwardly from the groove and rests againstan inner surface of the liner of the screw pump to reduce the slip flowfrom the high-pressure outlet to the low-pressure inlet. The ring sealand the groove are configured to retain the seal in the groove whileallowing radial displacement of the seal with respect to the thread asthe pump rotor is deflected.

In accordance with another embodiment disclosed herein, a twin-screwpump comprises a casing having an inlet and an outlet, a liner disposedinside of the casing and two rotors disposed inside of the liner. Eachrotor comprises a shaft, a thread disposed on a portion of an outersurface of the shaft, a groove on an outer surface of the thread and aring seal in the groove and configured to rotate with the shaft and toprotrude outwardly from the groove to rest against an inner surface ofthe liner. The ring seal and the groove are configured to retain theseal in the groove while allowing radial displacement of the seal withrespect to the thread as the pump rotor is deflected.

DRAWINGS

These and other features, aspects, and advantages of the presentinvention will become better understood when the following detaileddescription is read with reference to the accompanying drawings in whichlike characters represent like parts throughout the drawings.

FIG. 1 illustrates a conventional twin-screw pump.

FIG. 2 illustrates the pump slip flow path between rotor tips and theliner.

FIG. 3 illustrates a perspective view of a screw pump rotor inaccordance with aspects disclosed herein.

FIG. 4 illustrates rectangular cross-sectional ring seal in a groove ofa rotor tip.

FIG. 5 illustrates a close-up perspective view of a rotor tip inaccordance with an embodiment of the invention.

FIG. 6 illustrates a cross-sectional view of a rotor tip in accordancewith an embodiment of the invention.

FIG. 7 illustrates a cross-sectional view of a rotor tip in accordancewith another embodiment of the invention.

FIG. 8 illustrates a cross-sectional view of a rotor tip in accordancewith another embodiment of the invention.

FIG. 9 illustrates a cross-sectional view of a rotor tip in accordancewith another embodiment of the invention.

FIG. 10 illustrates a thread with groove having varying depth inaccordance with another embodiment of the invention.

DETAILED DESCRIPTION

Embodiments disclosed herein include screw pump rotors and ring sealsfor screw pump rotors. The pump rotor comprises a shaft, a thread on theshaft and a ring seal. The thread comprises a groove disposed on theouter surface of the thread and the ring seal is disposed in the groove.The groove and the seal are dimensioned to have a clearance between themto enable the seal to move radially with respect to the thread as thepump rotor is deflected. The seal and the groove are configured suchthat the seal is retained in the groove while allowing radialdisplacement of the seal with respect to the thread, as discussed inreference to FIGS. 3, 4 and 5. As used herein, the singular forms “a,”“an” and “the” include plural referents unless the context clearlydictates otherwise.

FIG. 3 illustrates a perspective view of an embodiment of a respectivescrew pump rotor 40. The threads 44 are helical and are over at least aportion of the shaft 42. The groove 48 is provided on the outer surfaceor tip 46 of the screw thread 44 that faces the pump liner. The ringseal 50 is typically spiraled into the thread 44 from a starting pointof the thread at an end of the shaft 42. The ring seal 50 is helical instructure and may have a length to cover any specific amount ofcircumferential displacement of the helical threads 44 of the rotor 40.In one embodiment, the ring seal 50 covers one complete revolution ofthe threads 44.

Pins 60 are used to hold the ring seal 50 in place inside and withrespect to the grooves 48 when the rotor is rotated. The pins 60 enablethe ring seal to rotate with the shaft. In one embodiment, the ringseals 50 are held in place by the pins 60 disposed once per revolution.In other embodiments, the pins 60 are disposed at any multiple orfraction of revolutions, depending on the circumferential length of thering seals 50.

As discussed above, the ring seal 50 is important to the overallperformance of a screw pump. A clearance between the rotor and the lineris required to allow for rotodynamic vibrations, manufacturingmisalignment, and rotor thermal expansion as well as for rotordeflection due to pressure. The ring seal 50 projects out from thethreads 44 and is configured to contact the pump liner, filling theclearance between the rotor and the liner. The ring seal 50 of theinstant invention provides improved sealing between the rotor and thepump liner. Improved sealing insures a reduction in solid particulateerosion or abrasion within clearances.

Occasionally, the ring seals wear down to an extent where they can getdislodged from the groove. For example, ring seals with rectangularcross-section, as shown in FIG. 4, do not have any constraint and arelikely to get dislodged from the groove after some wear which may leadto hard rubs or damage to the rotor or liner.

FIG. 5 illustrates a perspective view of a rotor tip, specificallyshowing a starting point of the thread 44. As explained above, the ringseal 50 is typically spiraled into the thread 44 from the starting pointof the thread.

In one embodiment, the groove 48 and the ring seal 50 comprise aninverted T-shape cross-section, as shown in FIGS. 5 and 6. The groove 48and the ring seal 50 are dimensioned to have a clearance between them.This clearance provides room for the ring seal 50 to move radially withrespect to the thread 44 as the pump rotor is deflected. The invertedT-shape cross-sections of the groove 48 and the ring seal 50 forms aninterlocking mechanism that limits the extent of radial displacement ofthe ring seal 50, as explained below.

The inverted T-shape cross-section of the ring seal 50 can be describedas having a first portion 82 substantially parallel to axis 80 of therotor and a second portion 84 substantially perpendicular to the rotoraxis 80 and the first portion 82. Similarly, the inverted T-shapecross-section of the groove 48 can be described as having a firstportion 86 substantially parallel to the axis of the rotor 80 and asecond portion 88 substantially perpendicular to the rotor axis 80 andthe first portion 86. The ring seal, when installed and under normaloperating conditions, is designed to spring outward to rest against oradjacent to an inner surface 52 of the pump liner 54, as best shown inFIG. 6. Typically, the second portion 84 of the ring seal contacts thepump liner 54. As the second portion 84 of the ring seal wears out, thering seal moves radially outward to maintain contact with the pump liner54. The first portion 82 of the ring seal is displaced radially outwardas the second portion 84 of the ring seal wears out. The second portion84 of ring seal can wear down to an extent until the first portion 86 ofthe groove prevents the first portion 82 of the ring seal from movingradially outward. Therefore, the T-shape cross-sections of the groove 48and the ring seal 50 facilitate in limiting the extent of radialdisplacement of the ring seal 50 and therefore prevent the ring seal 50from dislodging from the groove 48.

Prior to installation into the thread, the ring seal 50 has a freediameter. During installation of the ring seal 50 into the thread, thediameter of the ring seal is altered, which altered diameter is called afitted diameter. The contact pressure between the ring seal and theliner is affected by the difference between the free and the fitteddiameters of the ring seal. If the free diameter of the ring seal 50 islarger than the diameter of the liner 54, the ring seal 50 needs to becompressed during installation and the contact pressure will bemaintained at a higher level. If the free diameter of the ring seal 50is smaller than the diameter of the liner 54, the contact pressure isreduced or can be negligible until a combination of centrifugal forcesand pressure arise to deflect the ring seal 50 outward. Contact pressurewill generally decrease with the wear of the ring seal 50, extendinglife of the ring seal 50.

In operation, as best shown in FIG. 6, as the rotor turns, a profile ofincreasing pressure develops across the pump. The elimination orminimization of pump slip flow occurring between the rotor 40 and thepump liner 54 is accomplished by an outer surface of the ring seal 50being pushed against the inside surface 52 of the pump liner. Thespringing action of the ring seal 50 as well as a centrifugal load onthe ring seal 50 caused by the rotation of the rotor 40 pushes the ringseal against the inside surface 52 of the pump liner. A side surface 56of the ring seal 50 is also pushed against an inner surface 58 of thegroove 48 by the pressure difference from one side of the ring seal 50to the other. The clearance between the ring seal 50 and the groove 48enables slip flow 90 from beneath the ring seal 50, along the helicalgroove. Such slip flow 90 from beneath the ring seal flushes sand orother sediment that accumulates in the bottom of the groove 48.

The clearance between the ring seal 50 and the groove 40 enables slipflow 90 from beneath the ring seal 50 and is maintained to provide spaceunder the ring seal free of accumulation without allowing excessive slipflow. The slip flow 90 clears accumulation and allows the ring seal toretract, thereby reducing the contact pressure between the ring seal 50and the liner 54. Also, in one embodiment, the pins 60 disposed atmultiple or fractions of revolutions can be adapted to break the slipflow 90 from beneath the ring seals to control or limit the slip flow90.

As described earlier, the extent of radial displacement of the ring seal50 is limited. This form of mechanical restraint will only allow thering seal 50 to wear down to a point and the remaining part of the ringseal cannot escape the groove. The ring seal 50 is therefore alwaysretained in the groove 48.

In another embodiment as shown in FIG. 7, the ring seal 62 and thegroove 64 comprise a cross-section that is a mirror image of an L-shapecross-section. In another embodiment as shown in FIG. 8, the ring seal66 and the groove 68 are of dovetail-shape cross-section. In both theabove embodiments, the ring seals are spiraled into the grooves, theextent of radial displacement of the ring seals 62, 66 is limited, aclearance is maintained between the groove and the ring seal, and thering seals 62, 66 are prevented from getting dislodged from the grooves64, 68.

Another embodiment of the ring seal 70 is shown in FIG. 9. In thisembodiment, a low-pressure side 72 of the ring seal 70 and acorresponding side 74 of the groove 76 that is facing the low-pressureside 72 of the ring seal are inclined toward the high-pressure side 78of the ring seal 70. This configuration decreases the contact forcesbetween the ring seal 70 and the inner surface 52 of the pump liner 54,thereby reducing the wear rate of the ring seal 70.

The wear rate of the ring seal can also be reduced by sizing the grooveto drop the pressure below the ring seal so that the axial pressuredriven component that forces the ring seals against the liner 52, isminimized. The net outward pressure arises because high-pressure fluidleaking under the ring seal forces the ring seal outward. However,because the groove forms a continuous helix about the rotor, extendingthe groove to a low-pressure inlet and terminating or stopping thegroove before it connects to a high-pressure outlet of the pump canrelieve the outward pressure. The groove can be cut with varying depthto account for the integration of leakage about the several ring sealsand still provide suitable pressure relief, as shown in FIG. 10, wheredepth of the groove 48 decreases from a depth “d₁” to a depth “d₂”. Inone embodiment, the depth “d₁” to depth “d₂” decreases at a constantrate.

When the ring seals 50, 62, 66 and 70 are new, they can substantiallyseal the gap between rotor 40 and liner 54, even under full deflectionand continue to do so after some wear. At some point, the ring sealswear down to an extent where they cannot completely seal the gaps. Atthis point, the pump performance begins to wane. The performance can bemonitored and the worn out ring seals can be replaced with new ringseals. If the ring seals are not replaced at this stage, they mayeventually wear down to a point where they are flush with the rotor 40depending on the degree of eccentricity experienced between the rotorand the liner. If the rotor and liner are always concentric, theexcessive wear will be minimal. Although the ring seals provide minimalsealing benefit at this point, they stay in the groove and will notcause any problems with the system.

With respect to the above description, it should be realized that theoptimum dimensional relationships for the parts of the invention, toinclude variations in size, form function and manner of operation,assembly and use, are deemed readily apparent and obvious to thoseskilled in the art, and therefore, all relationships equivalent to thoseillustrated in the drawings and described in the specification areintended to be encompassed only by the scope of appended claims.

While only certain features of the invention have been illustrated anddescribed herein, many modifications and changes will occur to thoseskilled in the art. It is to be understood that the appended claims areintended to cover all such modifications and changes as fall within thetrue spirit of the invention.

1. A pump rotor for a screw pump, comprising: a shaft; a helical threadon the shaft, the thread comprising a groove disposed on an outersurface thereof, and a seal disposed in the groove, wherein the seal andthe groove are configured to retain the seal in the groove whileallowing radial displacement of the seal with respect to the thread asthe pump rotor is deflected.
 2. The pump rotor of claim 1, wherein theseal is a ring seal.
 3. The pump rotor of claim 2, wherein the ring sealis configured to protrude outwardly from the groove and to rest againstan inner surface of a liner of the screw pump.
 4. The pump rotor ofclaim 3, wherein the ring seal and the groove are configured to enableslip flows from beneath the ring seal.
 5. The pump rotor of claim 3,wherein the ring seal is configured to rotate with the shaft.
 6. Thepump rotor of claim 5, further comprising: a first and a second pinsdisposed in the groove, wherein the ring seal is disposed between thefirst and second pins.
 7. The pump rotor of claim 6, wherein the ringseal and the thread are helical.
 8. The pump rotor of claim 1, furthercomprising a plurality of pins disposed in the groove and the sealcomprises a plurality of ring seals, wherein each of the plurality ofring seals is disposed between a pair of the consecutive pins.
 9. Thepump rotor of claim 3, wherein the ring seal is a sacrificial wearcomponent of the pump rotor and the ring seal and the groove areconfigured to prevent the ring seal from getting dislodged from thegroove even after the ring seal is worn out.
 10. The pump rotor of claim3, wherein groove and the ring seal comprise an inverted T-shapecross-section.
 11. The pump rotor of claim 3, wherein ring seal and thegroove comprise a cross-section that is a mirror image of L-shapecross-section.
 12. The pump rotor of claim 3, wherein the ring seal andthe groove are of dovetail-shape cross-section.
 13. The pump rotor ofclaim 3, wherein a low pressure side of the ring seal and acorresponding side of the groove that is facing the low pressure side ofthe ring seal are configured to decrease the contact forces between thering seal and an inner surface of a liner of the screw pump.
 14. Amethod of reducing slip flow in a screw pump having a casing having alow-pressure inlet and a high-pressure outlet, a liner disposed insideof the casing, and a rotor disposed inside of the liner having a shaftand a thread disposed on an outer surface of the shaft, the methodcomprising: forming a groove on outer surface of the thread; anddisposing a ring seal in the groove such that the ring seal protrudesoutwardly from the groove and rests against an inner surface of theliner of the screw pump to reduce the slip flow from the high-pressureoutlet to the low-pressure inlet, the ring seal and the groove beingconfigured to retain the seal in the groove while allowing radialdisplacement of the seal with respect to the thread as the pump rotor isdeflected.
 15. The method of claim 14, wherein the ring seal and thegroove are configured to enable slip flows from beneath the ring seal.16. The method of claim 14, wherein the groove is formed varying indepth.
 17. The method of claim 14, further comprising enabling the ringseal to rotate as the shaft is rotated.
 18. A twin screw pump,comprising: a casing having an inlet and an outlet; a liner disposedinside of the casing; and at least two rotors disposed inside of theliner, each rotor comprising, a shaft; a thread disposed on a portion ofan outer surface of the shaft; a groove on an outer surface of thethread; and a ring seal in the groove and configured to rotate with theshaft and to protrude outwardly from the groove, wherein the ring sealand the groove are configured to retain the seal in the groove whileallowing radial displacement of the seal with respect to the thread asthe pump rotor is deflected.
 19. The pump of claim 18, wherein the ringseal and the groove are configured to enable slip flows from beneath thering seal.
 20. The pump of claim 18, wherein a low pressure side of thering seal and a corresponding side of the groove that is facing the lowpressure side of the ring seal are configured to decrease the contactforces between the ring seal and an inner surface of a liner of thescrew pump.
 21. A rotor for a screw pump comprising: a shaft; a helicalthread disposed on an outer surface of said shaft, said helical threaddefining a groove therein; a seal structure lockingly disposed withinsaid groove, wherein said seal and said groove are configured to allowradial displacement of the seal with respect to the thread as the rotoris deflected.